The present invention relates to an improvement in an impeller incorporated in a machine generally called turbomachinery such as a centrifugal pump for pumping liquid, or a blower or a compressor for pressurizing and delivering gas.
FIGS. 9A through 10B show a typical turbomachinery which is constructed by accommodating an impeller 6 having a hub 2, a shroud 4, and a plurality of blades 3 between the hub 2 and the shroud 4 in a casing (not shown in the drawings) having pipes and by coupling a rotating shaft 1 connected to a driving source to the impeller 6. In such an impeller, the blade tips 3a of the blades 3 are covered with a shroud surface 4a, and a flow passage is defined by two blades 3 in confrontation with each other, a hub surface 2a and the shroud surface 4a. 
When the impeller 6 is rotated about an axis of the rotating shaft 1 at an angular velocity xcfx89, fluid flowing into the flow passage from an impeller inlet 6a through a suction pipe is delivered toward an impeller exit 6b, and then discharged to the outside of the turbomachinery through a discharge pipe or the like. In this case, the surface facing the rotational direction of the blade 3 is the pressure surface 3b, and the opposite side of the pressure surface 3b is the suction surface 3c. 
The three-dimensional geometry of a closed type impeller as an example of impellers is schematically shown in FIGS. 9A through 10B in such a state that most part of the shroud surface is removed. In the case of an open type impeller, there is no independent part for forming the shroud surface 4, but a casing (not shown in the drawings) for enclosing the impeller 6 serves mechanically as the shroud surface 4. Therefore, there is no basic fluid dynamical difference between the open type impeller and the closed type impeller. Thus, only an example of the closed type impeller will be described below.
In the flow passages of such an impeller in a centrifugal turbomachinery, besides main flow flowing along the flow passages, secondary flows (flow having a velocity component perpendicular to that of the main flow) are generated by movement of low energy fluid in boundary layers on wall surfaces due to pressure gradients in the flow passages. The secondary flow affects the main flow intricately to form vortices or flow having non-uniform velocity in the flow passage, which in turn results in substantial fluid energy loss not only in the impeller but also in the diffuser or guide vanes downstream of the impeller. The total energy loss caused by the secondary flows is referred to as secondary flow loss. It is known that the low energy fluid in the boundary layers accumulated at a certain region in the flow passage due to the secondary flows causes a flow separation in a large scale, thus producing positively sloped characteristic curve and hence preventing the stable operation of the turbomachinery.
The secondary flow in the impeller is broadly classified into the blade-to-blade secondary flow generated along the shroud surface or the hub surface, and the meridional component of the secondary flow generated along the pressure surface or the suction surface of the blades. It is known that the blade-to-blade secondary flow can be minimized by making the blade profile to be backswept. Regarding the other type of the secondary flow, that is, the meridional component of the secondary flow, it is necessary to optimize the three-dimensional geometry of the flow passage, otherwise the meridional component of the secondary flow cannot be weakened or eliminated easily.
The mechanism of generation of the meridional component of the secondary flow is explained as follows: As shown in FIG. 9B, with regard to the relative flow in the flow passage, the reduced static pressure distribution, defined as p* (=pxe2x88x920.5xcfx81u2), is formed by the action of a centrifugal force W2/R based on the streamline curvature of the main flow and by the action of Coriolis force 2xcfx89Wxcex8 based on the rotation of the impeller, where W is the relative velocity of the flow, R is the radius of streamline curvature, xcfx89 is the angular velocity of the impeller, Wxcex8 is the component in the circumferential direction of W relative to the rotating shaft 1, p is the static pressure, xcfx81 is the density of fluid, u is the peripheral velocity at a certain radius from the rotating shaft 1.
The reduced static pressure p* has a distribution in which the pressure is high at the hub side and low at the shroud side, so that the pressure gradient balances the centrifugal force W2/R and the Coriolis force 2xcfx89Wxcex8 which are directed toward the hub side shown in FIG. 9B. In the boundary layer along the blade surface, since the relative velocity W is reduced by the influence of the wall surface, the centrifugal force W2/R and the Coriolis force 2xcfx89Wxcex8 which act on the fluid in the boundary layer become small. Accordingly, the centrifugal force and the Coriolis force cannot balance the reduced static pressure distribution p* of the main flow. As a result, the low energy fluid in the boundary layer flows towards an area of the low reduced static pressure p*, thus generating the meridional component of the secondary flow along the blade surface from the hub side toward the shroud side, on the pressure surface 3b or the suction surface 3c of the blade 3. In FIG. 9A, the meridional component of the secondary flow is shown by the dashed arrows on the pressure surface 3b of the blade 3 and the continuous arrows on the suction surface 3c of the blade 3.
The meridional component of the secondary flow is generated on both surfaces of the suction surface 3c and the pressure surface 3b of the blade 3. In general, since the boundary layer on the suction surface 3c is thicker than that on the pressure surface 3b, the secondary flow on the suction surface 3c has a greater influence on performance characteristics of a turbomachinery.
When the low energy fluid in the boundary layer moves from the hub side to the shroud side, fluid flow flowing from the shroud side toward the hub side is formed at the midpoint location between two blades to compensate for fluid flow rate which has moved. As a result, as shown schematically in FIG. 10A, a pair of vortices having a different swirl direction from each other is formed in the flow passage between two blades. These vortices are referred to as secondary vortices. Low energy fluid in the flow passage is accumulated due to these vortices at a certain location of the impeller where the reduced static pressure p* is low, and mixed with fluid which flows steadily in the flow passage, resulting in generation of great flow loss.
Furthermore, if the non-uniform flow generated by insufficient mixing of low energy fluid having a low relative velocity and high energy fluid having a high relative velocity is discharged to the downstream flow passage of the blades, then great flow loss is generated. Such a non-uniform flow leaving the impeller makes the velocity triangle unfavorable at the inlet of the diffuser and causes a separated flow on diffuser vanes or a reverse flow within a vaneless diffuser, resulting in substantial decrease of the overall performance of the turbomachinery.
Therefore, as shown in FIGS. 11A and 11B, in order to optimize the distribution of the reduced static pressure p* in the impeller, it is considered to design the impeller as follows: The blade is leaned toward a circumferential direction, between the location of non-dimensional meridional distance m=0 (impeller inlet) and the location of non-dimensional meridional distance m=1.0 (impeller exit), so that the blade at the hub side precedes the blade at the shroud side in a rotational direction of the impeller. Further, the blade lean angle, defined as an angle between a surface perpendicular to the hub surface and the blade centerline on the cross-sectional view of the flow passage in the impeller, shows a decreasing tendency as the non-dimensional meridional distance m increases.
According to the impeller having the above structure, since the blade is leaned toward a circumferential direction so that the blade at the hub side precedes the blade at the shroud side in a rotational direction of the impeller, a force having a component toward the shroud surface 4 acts on the fluid, the reduced static pressure p* in the flow passage has a higher value at the shroud surface and a lower value at the hub surface 2 to balance the component of the force toward the shroud surface. Further, since the blade lean angle shows a decreasing tendency as the non-dimensional meridional distance m increases, the effect of the blade lean is higher than that in the case where the blade at the shroud side is leaned toward the circumferential direction.
However, in the conventional technology having the above structure, as shown in FIG. 11A, since an angle between a line connecting the center of the blade at the shroud side and the center of the blade at the hub side and a surface perpendicular to the hub surface as viewed from the direction of the impeller exit (rake angle xcex3) is extremely large, the blade is deformed by the rotation of the impeller so as to be raised, causing a large bending stress at the blade base.
Further, as shown in FIGS. 11A and 11B, at the impeller inlet, since an angle between a line connecting the center of the blade at the shroud side and the center of the blade at the hub side and a line connecting the center of the blade at the hub side and the center of the impeller (lean angle xcex4) is formed, the blade is deformed by the rotation of the impeller so as to be raised, causing a large bending stress at the blade base. Further, in the case of a closed type impeller having a cover at the shroud side of the impeller, complicated stresses are caused at various portions of the blade due to formation of the lean angle and the rake angle.
In the case where the impeller is manufactured by welding, the blade base is a part of the welded structure. Accordingly, insufficient welding tends to be caused by the leaned blades, initiating cracks on the welded portion due to rotation and causing a breakdown. Further, since the large stress at the blade base affects the useful life of the impeller, a high degree of welding technology and a high-quality material are required to thus raise manufacturing cost. In the case where the blades are manufactured by mechanical cutting, complicated working is required for mechanical cutting to thus raise manufacturing cost.
The present invention has been made in view of the above drawbacks. It is therefore an object of the present invention to provide a centrifugal turbomachinery having a good performance which can effectively reduce the secondary flow in the flow passage of the impeller and minimize the loss caused by the secondary flow without an excessive increase in manufacturing cost.
According to a first aspect of the present invention, there is provided an impeller having a plurality of blades between an inlet at a central portion and an exit at a peripheral portion, and a flow passage formed between the blades for delivering fluid from the inlet to the exit by rotation of the impeller, characterized in that: the blade is leaned toward a circumferential direction so that the blade at the hub side precedes the blade at the shroud side in a rotational direction of the impeller at an exit side; a blade lean angle, defined as an angle between the blade and a surface perpendicular to a hub surface as viewed from the direction of the exit of the flow passage, shows a decreasing tendency from the inlet to the exit; and a blade centerline at the hub side and a blade centerline at the shroud side as viewed from the front direction at the inlet intersect at a point where non-dimensional radius location, defined as a ratio of the radius of the intersection to the radius of the exit, ranges from 0.8 to 0.95.
According to another aspect of the present invention, there is provided a turbomachinery having a rotatable impeller incorporated in a casing, the impeller having a plurality of blades between an inlet at a central portion and an exit at a peripheral portion, and a flow passage formed between the blades for delivering fluid from the inlet to the exit by rotation of the impeller, characterized in that: the blade is leaned toward a circumferential direction so that the blade at the hub side precedes the blade at the shroud side in a rotational direction of the impeller at an exit side; a blade lean angle, defined as an angle between the blade and a surface perpendicular to a hub surface as viewed from the direction of the exit of the flow passage, shows a decreasing tendency from the inlet to the exit; and a blade centerline at the hub side and a blade centerline at the shroud side as viewed from the front direction at the inlet intersect at a point where non-dimensional radius location, defined as a ratio of the radius of the intersection to the radius of the impeller exit, ranges from 0.8 to 0.95.
The hub, the shroud, and the blade may be integrally formed of metal.